The present invention generally relates to a heat exchanger and more particularly, to a fin-tube type heat exchanger to be employed in air conditioning, refrigeration and cold storage units, etc., for facilitating heat transfer between a cooling medium and a fluid such as air or the like.
Conventionally, as shown in FIG. 5, the fin-tube type heat exchanger of the above-described type is constituted by many plate fins 1 arranged in a parallel relation to each other at predetermined intervals, and heat exchanging tubes 3 extending through said plate fins 1 in a direction at right angles thereto. An air stream A is caused to flow between the plate fins 1 for undergoing heat exchange with the cooling medium flowing within the heat exchanging tubes 3. In recent years, although a reduction in size and higher performance have been required for such a fin-tube type heat exchanger, due to the fact that the air velocity between the plate fins is suppressed to reduce noises, etc., the heat resistance offered at the air side is high compared to that offered within the heat exchanging tubes. Therefore, at present, to reduce the difference in heat resistance offered at the air side and within the heat exchanging tubes, the heat transfer area at the air side is enlarged. However, since the expansion of the heat transfer area is limited by physical restraints and economics and by the desirability to save space, etc., a reduction in the heat resistance offered at the air side has been an important characteristic to be achieved in the fin-tube type heat exchanger of this kind.
In FIGS. 6 and 7, there is shown one example of a conventional fin-tube type heat exchanger in which fin collars 2 are erected on a plate fin 1 at equal intervals. Between said fin collars 2, cut and raised portions 1a are formed so as to be open to air stream A only at the side of the plate fin 1 from which the fin collars 2 protrude and so as to project from the surface of the base plate of the plate fin 1 by distances equal to each other. The cut and raised portions referred to above are intended to prevent the development of a thermal boundary layer. The heat exchanging tubes 3 are so arranged that a pitch L.sub.1 ' over which the tube rows are spaced in the direction of the air stream A is set at 1.9 to 2.2 times the outer diameter Do' of said tubes 3, while a pitch L.sub.2 ' over which the tubes are spaced in each row in the direction perpendicular to the air stream A is set at 2.2 to 2.5 times the outer diameter Do' of said tubes 3. The tubes 3 extend through the plate fin 1 in close contact with inner surfaces of the fin collars 2. The above heat exchanging tubes 3 have a U-shape, with opposite ends thereof being connected by bends (not particularly shown). In FIG. 6, numerals 4a and 4b represent dead air regions formed at slip stream sides of the heat exchanging tubes 3. In the known construction as described above, however, an optimum tube arrangement for maximizing the overall heat transfer coefficient at the air side, based on the same fan power standard by taking into account the flow resistance .DELTA.P of the air stream, is not realized, thus resulting in an uneconomical design. Moreover, since the cut and raised portions 1a do not extend from the base plate portion in a direction perpendicular to the air stream A flowing between the tubes 3, the average heat transfer distance from front and rear portions of said tube 3 to the cut and raised portions 1a tends to be long, with a consequent lowering of the fin heat transfer efficiency. And, a sufficient boundary layer leading edge effect is not produced since each cut and raised portion 1a has a short leading edge. Furthermore, due to the leg portions of the cut and raised portions 1a being superposed in a direction normal to the leading edge of the plate fin la, the air stream A is not altered in direction even after passing through the cut and raised portions 1a, thus making it impossible to accelerate the generation of turbulent flow. Meanwhile, dead air regions 4a and 4b are relatively large, resulting in a corresponding reduction in the effective heat transfer area. Additionally, since the neighboring cut and raised portions 1a are of the same length, the leg portions thereof are undesirably superposed as viewed in the direction of flow of the air stream A and thus, the resistance against flow is concentrated resulting in a non-uniform flow rate distribution, whereby the effect of the cut and raised portions 1a cannot be fully utilized.